U.S. Pat. No. 3,583,293 refers to numerous ways in which a piston for use in an internal combustion engine can be designed for optimum efficiency.
Proper functioning of the piston and piston rings is essential for an efficient compression-combustion-expansion process, and therefore the requirements made on these components are many and severe.
There are two different types of piston rings in normal internal combustion engines, compression rings and oil scraper rings. The most important requirement of the compression rings is to form a gas-tight seal between the piston and the cylinder wall. If the compression ring seal is ineffective, excessive "blow-by" will occur, resulting in an overall decrease of engine efficiency and life. The oil scraper ring acts to prevent the drawing of oil into the area above the sealing rings by applying significant pressure, known as "radial load", against the cylinder wall.
The basic design of the piston ring has not changed since its original conception and there are certain drawbacks to the conventional piston ring arrangement that should be examined in order to better understand how the present invention overcomes the shortcomings of current piston and ring design.
Piston rings currently in use seal against gas pressure created above the piston during the compression-combustion-expansion process by two different methods; one is the bearing of the ring against the walls of the cylinder (i.e., radial loading) and the second is the bearing of the ring against the upper or lower face of the ring land.
With respect to the former, effective sealing by the piston ring can only be accomplished when there is adequate gas pressure behind the ring forcing it radially outward. The radial force exerted by a piston ring in tension is not in itself enough to effect proper sealing under most conditions. The pressure behind the ring must be nearly equal to the pressure above the piston, otherwise the ring will collapse radially inward, and no longer seal. Indeed, during normal operation of an engine with rings, grooves, and cylinders in good condition, the pressure behind the topmost piston ring is nearly equal to the pressure in the cylinder and progressively lower pressures exist behind the second and third rings.
Thus, the compressed gases must have free access to the space behind the rings in order to have effective sealing. Ring-land to ring clearance must provide the necessary space for gas accessibility, but must not allow the ring to slam up and down in the groove. Such motion will lead to an ineffective seal at the top or bottom of the ring land and lead to excessive wear of this area of the piston. In the case of aluminum/alloy pistons, this motion can also cause the ring land to break away from the piston.
In addition, "ring flutter", which is encountered at high RPM, is caused when the inertial load of the ring during the compression stroke exceeds the combined gas pressure and wall friction which seats the ring in the ring-land bottom, causing the ring to lift off its seat. Once this occurs, the gases behind the ring, which are essential to the sealing ability of the ring, escape and the ring collapses.
In summary, the most effective seal obtainable from a traditional piston/ring configuration relies on the optimum and equivalent translation of combustion pressure above the piston to outward pressure of the ring against the cylinder bore, a relationship which must occur while maintaining an effective seat between the ring and the top or bottom of the ring groove.
It is generally accepted that the piston and piston ring assembly are the most critical components and the highest contributors to friction in the internal combustion engine.
In internal combustion engines, friction at the piston assembly occurs in several ways. First, the angularity of the rod in relation to the cylinder bore creates a side load against the wall. The amount of friction resulting from this effect is determined by the length of the rod (its overall angularity) and the design and integrity of the piston skirt, which, in traditional designs, can have a tendency to radically deflect or deform in response to side thrusts created by the rod.
Second, the radial pressure of the ring against the cylinder wall, which is caused by the necessary presence of compressed gas behind the rings, as well as the inherent radial loading of the ring, cause the piston rings to drag against the bore. This is particularly true of rings which are spring loaded and have a large axial and/or radial dimension, such as the oil scraper ring. However, generally speaking, the more rings that are present, the greater the friction generated.
U.S. Pat. No. 3,583,293 presents the inventor's design of a piston with greater structural rigidity. In this design, the stability of the piston in the bore, maintained previously to a great extent by an oil ring with a very high radial load, was instead maintained by a very rigid piston with a skirt having multiple parabolas which held the deflection of the piston in the bore to a minimum. The enhanced stability offered by this design was further complemented by the presence of the single ring land, which allowed placement of the wrist pin hole closer to the crown, increasing connecting rod length and decreasing the angularity of the rod.
The single, stepped ring groove housed three component parts. The lower ring, situated in the stepped portion of the groove, was called a "rail", and was of diminutive radial and axial dimension. It acted to close the end gap of the upper compression ring, providing a better seal. Above the rail was a more standard compression ring backed by an oil ring expander. Placement of the oil expander ring behind the compression ring was intended to achieve the proper radial load of the compression ring against the cylinder bore.
In addition, the design provided for numerous perforations, located at the bottom edge of the ring belt, called "oil drainbacks", which diverted oil on the cylinder bore through the piston wall, reducing the accumulation of pressure at the bottom of the lower ring created by the pistons' downward travel. This reduced the radial load requirements of the ring package, allowing better contact between the entire ring package and the cylinder wall. The diversion of oil to the underside of the pistons' crown, combined with the improved heat dissipation characteristics of the less massive ring belt, also reduced the operating temperature of the piston.
In practice, the rigid skirt with its oil drainbacks, combined with this innovative piston ring configuration, led to a significant reduction in overall friction and provided for a better seal between the rings and the cylinder bore. However, it soon became apparent that the improved seal resulted in less effective oil control. Somehow, oil was travelling beyond the piston into the combustion chamber. Initially, it was not clear whether the oil was passing beyond the face of the ring package along the cylinder wall or behind the rings at the inner portion of the ring groove. In an effort to determine the cause, the standard oil ring expander behind the compression ring was replaced by a round (in cross section) elastomer "O" ring. The purpose here was to inhibit the flow of oil around the back of the rings, eliminating this route as a variable.
Subsequent testing indicated that oil consumption was radically increased. At this juncture, it was hypothesized that the oil was bypassing the face of the ring package during the intake stroke in response to an increased vacuum created above the piston as it travelled downward in the cylinder bore. This phenomenon, caused by the pressure differential, was made only more acute with the use of the "O" ring behind the top compression ring. In effect, the oil was being "pushed" upward, making its way past the piston by "hydroplaning" the ring package off the cylinder wall during the intake cycle. While virtually eliminating blow-by on the compression stroke, the new ring pack configuration exacerbated the oil control problems of the earlier design by preventing any measure of pressure equalization during the intake stroke.
In order to maintain the benefits of an improved seal during the compression stroke without the deleterious oil control effects, it was determined that the pressure differential created above and below the piston during the intake stroke had to be eliminated. As discussed in U.S. Pat. No. 3,839,996, a vacuum balance system (VBS) was, therefore, implemented through which the combustion chamber and crankcase were connected, sharing a common environment during the intake cycle. The addition of VBS to the rigid skirt, single ring combination, virtually eliminated oil consumption while retaining the benefits of a lighter, more rigid piston assembly with improved sealing capabilities.